Dry screw driver

ABSTRACT

An embodiment of a dry screw compressor has a male rotor at a peripheral speed lower than approximately 80 m/s. The compressor includes: a casing body having an inlet for a gaseous fluid to be taken in and at least an outlet for the compressed fluid; at least a male rotor and at least a female rotor meshed together, said rotors being arranged inside said casing body whereby the ratio between the length and the external diameter of the male rotor is higher than or equal to approximately two, and the winding angle of the male rotor is smaller than or equal to approximately 300°.

PRIORITY CLAIM

The present application is a national phase application filed pursuant to 35 USC §371 of International Patent Application Serial No. PCT/IB2010/001706, filed

July 9, 2010; which further claims the benefit of Italian Patent Application Serial No. PR2009A000054 filed Jul. 10, 2009; all of the foregoing applications are incorporated herein by reference in their entireties.

TECHNICAL FIELD

An embodiment relates to a dry screw compressor for a gas, in particular air, for use in pressure applications (e.g. in the conveyance of granulates or powders, or in water treatment, where large amounts of air must be conveyed to start and help aerobic reactions) and in vacuum applications (e.g. in gas, fume or steam exhaust systems). In particular, an embodiment of the present dry screw compressor is used in applications with low differential pressures comprised between 1 bar and 3 bars and under vacuum to a threshold absolute pressure of 150 mbars.

BACKGROUND

As already known, low differential pressure applications (lower than 1 bar) use lobe compressors. These are compressors wherein two lobe rotors (usually two or three lobes) with parallel axes mesh together and synchronically rotate in opposite directions.

However, these lobe compressors, although being structurally simple, economical and able to guarantee a good flow, have a scarce thermodynamic efficiency.

Therefore, it has been projected a screw compressor which could work under low pressure, with a high flow and a thermodynamic efficiency characteristic of an internal compression machine, but whose structural characteristics were as similar as possible to a lobe compressor.

As known, a normal screw compressor under high pressures includes at least a male rotor and at least a female rotor meshing together during the rotation around respective axes and housed within a casing body. Each of the two rotors has screw-shaped ribs that mesh with corresponding screw-shaped grooves of the other rotor. Both the male and female rotor show, in cross section, a predetermined number of teeth corresponding to their ribs and a predetermined number of valleys corresponding to their grooves. The casing body has an inlet for the gas to be taken in and an outlet (also called “delivery outlet”) for the compressed gas. The intake gas is compressed between the two moving rotors and arrives to the outlet under the requested pressure.

Furthermore, it is known that dry screw compressors, generally indicated as “oil-free”, as opposed to oil injection compressors, are largely used in applications wherein the level of contaminants must be kept below a determined percentage threshold (usually very low).

In recent years, some manufacturers have proposed dry screw compressors for differential pressures approximately between 3 and 10 bars, thus re-adjusting the technology of oil injection screw compressors for applications under high pressures (higher than approximately 10 bars).

However, the manufacture of such dry screw compressors is quite sophisticated and expensive, since it typically must take into account the remarkable mechanical and thermal stresses to which the rotors are subjected. In particular, in order to avoid an excessive bending under load, the ratio between the length and the external diameter of the male rotor is usually approximately between 1.5 and 1.8; this ratio may strongly limit the compressor capacity and may require the insertion in the compressor structure of a gear multiplier in order to start the rotors at very high peripheral speeds, usually approximately >150 m/s.

By altering the delivery outlets, the aforesaid compressors can be also used under differential pressures approximately between 1 bar and 3 bars. However, a drawback of these low pressure compressors is represented by their having the same structural complexity of high-pressure compressors.

SUMMARY

Therefore, an embodiment is a dry screw compressor which can work under low pressure, with a high flow and a thermodynamic efficiency typical of these kinds of machines.

In particular, an embodiment is a dry screw compressor under low differential pressures (approximately between 1 and 3 bars) and with a high flow, which is structurally simple, economical and easy to maintain.

Moreover, a further embodiment is a dry screw compressor which is also suitable for under vacuum applications, up to a threshold of approximately 150 mbar of absolute pressure.

BRIEF DESCRIPTION OF THE DRAWINGS

Characteristics and advantages of one or more embodiments will become more apparent from the following approximate, and hence non-restrictive, description of one or more embodiments of a dry screw compressor, as illustrated in the appended drawings, wherein:

FIG. 1 illustrates a longitudinal cross section of a dry screw compressor according to an embodiment;

FIG. 2 illustrates a three-dimensional view, in longitudinal section, of some details belonging to an embodiment of the dry screw compressor belonging in turn to the dry screw compressor of FIG. 1;

FIG. 3 illustrates a cross section (not in scale) of an embodiment of the rotors used in the compressor according to FIGS. 1, 2; and

FIG. 4 illustrates a three-dimensional schematic side view (not in scale) of a male rotor used in the dry screw compressor according to an embodiment.

DETAILED DESCRIPTION

With reference to the alleged FIGS., 1 indicates a dry screw compressor for gas, in particular air, according to an embodiment.

The compressor 1 can be used both under pressure and under vacuum.

The compressor 1 includes at least a male rotor 2 and at least a female rotor 3, meshed together (FIGS. 1, 2, 3).

The embodiment here described and illustrated provides for a single male rotor 2 and a single female rotor 3 housed within a single casing body 4.

In particular, this casing body 4 is obtained by coupling two communicating cylinders (not shown) so that they define a single cavity 5 housing the rotors 2, 3.

An alternative embodiment (not shown) provides for a plurality of conjugated pairs of male rotors 2 and female rotors 3.

As shown in FIG. 1, the female rotor 3 is keyed on a shaft 17 (having an axis of rotation (01)), whereas the male rotor 2 is keyed on a shaft 18 (having an axis of rotation (02)). In particular, the first axis of rotation (01) is arranged at a certain distance (I) (generally known as “center distance”) from the second axis of rotation (02). The first axis of rotation (01) and the second axis of rotation (02) are parallel to each other.

Each of said rotors 2, 3 has screw-shaped ribs meshing with screw-shaped grooves formed between corresponding screw-shaped ribs of the other rotor 2, 3. In this way, in cross section (FIG. 3), the male rotor 2 shows lobes 6 (or teeth) and valleys 7 meshing with corresponding valleys 8 and lobes 9 (or teeth) of the female rotor 3.

Moreover, FIG. 3 shows some main dimensional parameters characterizing the profiles of the rotors 2, 3. In particular, it can be seen an external circumference (Cef) of the female rotor 3 and an external circumference (Cem) of the male rotor 2.

Moreover, as shown in FIG. 1, the length (Lm) of the male rotor 2 corresponds to the length (Lf) of the female rotor 3.

Conjugated profiles identical to those shown in the present FIG. 3 have been described and claimed in the international patent application PCT/IB2010/051416, which is incorporated by reference, and whose content is part of the present detailed description since, in combination with the geometrical elements described hereinafter, it allows to maximize the compressor capacity and to minimize the gas leaks normally occurring in the coupling areas between the rotors, and between the rotors and their casings.

In fact, an embodiment and with a particular reference to FIG. 4, the “winding angle” (φ) is formed by the angle of a generic helix 40 (described by the head of a generic tooth) between a segment OA, coupling the axis (02) of the male rotor 2 to the helix 40 on a first end plane (π1) of the rotor 2, and a segment O′B′, also coupling the axis (02) to the helix 40 on a second end plane (π2) of the rotor 2 opposed to the first end plane (π1).

As also shown in FIG. 4, the rotor 2 includes three helixes 30, 40, 50, parallel to each other, described by the heads of the relative teeth.

Furthermore, the term “length (Lm)” of the male rotor 2 defines the distance between the two end planes (π1), (π2), the term “pitch (Pz)” between two helixes 30, 40 defines the distance between point B and point B1, and the term “angle of the helix” (ψ) defines the angle between the tangent (r) to the helix 40 in any point (P) and the axis (02) of the male rotor 2.

It has been found that the ratio between the length (Lm) and the external diameter (Dm) of the male rotor 2 (see also FIG. 4), in an embodiment, must be higher than or equal to approximately 2, to maximize the compressor capacity and, therefore, together with the conjugated profiles of the lobes of the rotors, to guarantee high gas flows. Said (Lm)/(Dm) ratio may be approximately between 2 and 3. In this context, external diameter (Dm) means the diameter of the external circumference (Gem) of the male rotor 2 (FIG. 3).

Moreover, it has been found that, in an embodiment, in order to maximize the compressor capacity, if the other geometric dimensions are equal, the maximum value of the winding angle (φ) must be approximately 300°; in fact, by increasing the value of the winding angle (φ), and with an equal length (Lm), an equal diameter (Dm) and an equal profile of the tooth of the male rotor 2, the overlap between the teeth of the two rotors 2, 3 consequently increases, with a following reduction of the total capacity of the compressor 1.

Moreover, the values (Lm), (Pz) and the angles (φ), (ψ) are geometrically related to each other.

Therefore, in an embodiment, it is possible to projectually determine the optimal values of parameters (Lm), (Dm), (Pz), (ψ) in order to define an optimal value of the “winding angle” (φ) giving the maximum gas flow at reduced peripheral speed of the male rotor 2 and under reduced pressure.

The number of lobes 6 of the male rotor 2 may be different from the number of lobes 9 of the female rotor 3. In particular, the number of lobes 6 of the male rotor 2 may be lower than the number of lobes 9 of the female rotor 3 by at least one unity. For example, in the embodiment here described and illustrated, the number of lobes 6 of the male rotor 2 corresponds to three, whereas the number of lobes 9 of the female rotor 3 corresponds to five. In another embodiment (not shown), the number of lobes 6 of the male rotor 2 corresponds to four, whereas the number of lobes 9 of the female rotor 3 corresponds to six.

The two rotors 2, 3 are kept in the reciprocal position by means of the synchronization gear formed by two toothed wheels 20 a and 20 b of the known kind (FIG. 1).

In order to allow a correct working of the compressor 1 in an embodiment, the transmission ratio between the synchronization gears 20 a, 20 b is equal to the ratio existing between the number of teeth of the two rotors 2, 3.

The driving shaft is the shaft 17 on which the female rotor 3 is keyed because it is the one with more teeth, so that each rotation of this shaft 17 corresponds to the filling of a larger number of gaps and, in short, to a larger volume conveyed by the compressor 1.

As shown in more detail in FIG. 2, the casing body 4 has an inlet 10 for a gaseous fluid to be taken in, flowing according to an arrow (F1), and at least an outlet 11 (or delivery outlet) for the compressed fluid flowing according to an arrow (F2). Said outlet 11 defines an opening 12 formed in the casing body 4.

The compressor 1 uses bearings of a known kind. In particular, the radial loads are sustained by a first group 19 a of radial ball bearings arranged close to the inlet 10 and by a second group 19 b of cylindrical ball bearings arranged close to the outlet 11. The axial loads, on the other hand, are sustained by a third group 19 c of oblique contact ball bearings arranged beside the bearings of the second group 19 b.

In the particular embodiment shown in FIG. 1, the compressor 1 is provided with an electric motor 16 whose rotor is keyed on the shaft 17 of the female rotor 3 for starting its rotation around the first axis of rotation (01). The motor 16 may be a permanent magnet motor. Such a permanent magnet motor 16 may be of the kind cooled by water circulation. As an alternative, a permanent magnet motor of the air-cooled kind may be used.

As previously stated, the motor 16 may be keyed on the shaft 17 of the female rotor 3, namely it may be aligned with said shaft 17.

When no speed variation of rotors 2, 3 is required, the compressor 1 may be coupled to an electric motor (not shown) by means of a “belt and pulley” drive (not shown).

The operation of the dry screw compressor according to an embodiment is described hereinafter.

The gas (e.g. air) is taken in by the compressor 1 and, through the inlet 10, enters into the casing body 4 (FIGS. 1, 2). During the rotation, the screw-shaped ribs of the female rotor 3 mesh together with the screw-shaped grooves of the male rotor 2, and vice versa. In embodiments with no contact between the rotors 2, 3, the correct transmission/multiplication ratio between the rotors 2, 3 is actuated by means of the synchronization gears 20 a, 20 b.

By longitudinally crossing the casing body 4, the gas is compressed between the “coils” of the two rotating rotors 2, 3, thus reaching the outlet 11.

An embodiment wherein the opening 12 is arranged on the side surface of the casing body 4, is used for “intermediate” compression ratios R, e.g. approximately between 1 and 4; whereas, in another embodiment, the opening 12 is arranged in correspondence to an end of the casing body 4 (on the plane (π1); see FIG. 1); this latter embodiment solution is chosen for “high” compression ratios (R), e.g. approximately between 4 and 10. Both embodiments may be provided with shaping means (not shown) defining the actual dimension of opening 12 corresponding to the desired compression ratio (R).

The aforesaid description clearly shows the characteristics of the dry screw compressor according to one or more embodiments, as well as its advantages.

In particular, the ratio between the length and the external diameter of the male rotor (higher than or equal to approximately two) is made possible by low differential pressures (approximately between 1 bar and 3 bars) or by the threshold absolute pressure of approximately 150 mbars for under vacuum applications.

Moreover, the choice of the profile geometry and the operation of the compressor by means of the shaft of the female rotor allow to maximize the compressor capacity, with rotors of the same length, thus allowing to reach the requested high flow at a peripheral speed of the male rotor 2 lower than approximately 80 m/s.

Furthermore, the geometry of the profiles of the two coupled rotors allows obtaining a shorter contact line between the rotors with a better seal, thus reducing the blow by.

Moreover, thanks to the fact that the compressor works at peripheral speeds of the male rotor lower than approximately 80 m/s, the peripheral speed of the female rotor is even lower, and, therefore, the rotor of the electric motor can be directly keyed on the shaft of the female rotor (namely with no interposition of multiplying gears), thus obtaining a compressor which is structurally simple, compact and having a higher energetic efficiency. This makes use of the multiplying ratio of synchronization gears of the rotors, corresponding to the ratio between the number of lobes of the female rotor and the number of lobes of the male rotor (in the described embodiment, it corresponds to 5/3=1.66667). This avoids the utilization of toothed-wheel multipliers integrated in the compressor, with a resulting advantage in structural simplicity, encumbrance, cost and noise.

Furthermore, the energetic efficiency of the compressor is also provided by the use of a permanent magnet motor, characterized by low consumptions over a large range of speeds. In particular, this kind of permanent magnet motor has higher efficiencies than the three-phase asynchronous electric motor used in the known art, especially at reduced speeds. Among other things, the use of a water-cooled permanent magnet motor allows a reduction in size and weight of the motor, thus allowing its direct arrangement of the shaft of the female rotor, exploiting the radial bearings of the compressor.

Finally, the optimization of the energetic efficiency is also obtained thanks to the use of a delivery outlet whose size varies according to the desired compression ratio, thus producing an extremely versatile and modular compressor.

From the foregoing it will be appreciated that, although specific embodiments have been described herein for purposes of illustration, various modifications may be made without deviating from the spirit and scope of the disclosure. Furthermore, where an alternative is disclosed for a particular embodiment, this alternative may also apply to other embodiments even if not specifically stated. 

1. A dry screw compressor having a male rotor whose peripheral speed is lower than approximately 80 m/s, the compressor comprising: a casing body having an inlet for a gaseous fluid to be taken in and at least an outlet for the compressed fluid; at least a male rotor and at least a female rotor meshed together, said rotors being arranged inside said casing body; wherein the ratio between the length and the external diameter of the male rotor is higher than or equal to approximately two, and the winding angle of the male rotor is smaller than or equal to approximately 300°.
 2. The compressor according to claim 1, wherein the driving shaft is a shaft on which said female rotor is keyed.
 3. The compressor according to claim 2, further comprising an electric motor operationally acting on the shaft of the female rotor for starting its rotation around a first axis of rotation.
 4. The compressor according to claim 3, wherein in the rotor of said electric motor is keyed on said shaft of the female rotor.
 5. The compressor according to claim 3, wherein said electric motor is a permanent magnet motor.
 6. The compressor according to 1, wherein said outlet defines an opening formed in the casing body, the actual size of said opening being variable by means of shaping means in order to obtain a predetermined compression ratio.
 7. The compressor according to claim 1, wherein the compressor can be used in applications under differential pressures approximately between 1 bar and 3 bars.
 8. The compressor according to claim 1, wherein the compressor can be used in applications under vacuum up to a threshold absolute pressure of approximately 150 mbars. 